Heat Exchanger with a Tube Bundle and Shell with a Flow at the Shell Side with Improved Efficiency

ABSTRACT

There is described a heat exchanger comprising a plurality of tubes arranged parallel with each other in order to form one or more tube bundles axially inserted into a shell. A first fluid, fed through one or more first inlet nozzles flows inside the tubes and a second fluid, fed through at least one second inlet nozzle, flows inside the shell in order to perform the heat exchange with the first fluid through the walls of the tubes. Inside the shell two or more baffles are formed and arranged perpendicularly with respect to the centre axis of the shell. Between each baffle and the inner walls of the shell there is defined a corresponding window constituted by the cross-section for the passage of the second fluid, with a crossing direction parallel with the centre axis of the shell and the tubes, delimited by the free edge of the respective baffle on the one side and by the inner profile of the shell, at the intersection of the shell with the plane of the baffle, on the other side. The cross-section for the passage of the second fluid placed between two adjacent baffles is constant and has a rectangular shape, that is to say, with all the inner angles congruent with each other. Each baffle has a rectangular shape. Each window has a rectangular shape and has no tubes therein.

CROSS-REFERENCE TO RELATED PATENT APPLICATIONS

This application is a continuation of International application number PCT/EP2016/081638 filed on Dec. 19, 2016.

This patent application claims the benefit of International application No. PCT/EP2016/081638 filed on Dec. 19, 2016 and Italian application No. 202015000089247 of Dec. 30, 2015, the teachings and disclosure of which are hereby incorporated in their entirety by reference thereto.

BACKGROUND OF THE INVENTION

The present invention generally relates to a heat exchanger with a tube bundle and shell and, more particularly, to the geometric shape of three characteristic elements of the heat exchangers with a tube bundle and shell, that is to say, the geometric shape of the inner portion of the shell, in which the passage of the fluid at the shell side is brought about, the geometric shape of the transverse plate type baffles and the geometric shape of the so-called “window” in the case of the particular configuration in which there is no provision for the presence of tubes which extend in the window itself (so-called “no tubes in window” configuration).

A heat exchanger with a tube bundle and shell is an apparatus having the main function of transmitting the heat between two fluids at different temperatures. This heat exchanger is constituted by a tube bundle which is usually produced from a material with a high level of thermal conductivity and which is positioned inside a receiver which is generally of cylindrical form and which is called a shell. The tube bundle may have a single passage or multiple passages at the tube side. Both cases can be produced using straight tubes, the open ends of which are fixed to two perforated plates which are called tube plates (designated 34 in the appended Figures), by means of welding or mechanical expansion (rolling expansion). The particular geometry of the closure heads is what will produce the differentiation between the single passage and the multiple passages. However, a double-passage configuration at the tube side most used is produced by using so-called “hairpin” or “U”-shaped tubes (non-straight), in which both the open ends of each tube are fixed to the same tube plate. The heat exchange is brought about between the two fluids which have different temperatures, without there being any mixing between them: the first fluid flows inside the tubes (“tube side”) and is kept physically separated (no direct contact but only indirect contact by means of the tube exchange surface) from the second fluid which instead flows at the outer side of the tubes (“shell side”). There are provided inside the shell one or more baffles which serve to support the tubes and to control the main flow direction of the fluid which flows in the shell.

FIG. 1A shows the main components of a heat exchanger 10 with a tube bundle 12 and a shell 14, with reference for the sake of simplicity to the specific configuration with a single fluid passage at the shell side and a single fluid passage at the tube side. The fluid at the tube side is introduced into the heat exchanger 10 via a first inlet nozzle 16, flows back into a distribution chamber 18 which is formed inside the front head 20 and is generally distributed in equal quantities in each tube, thereby forming parallel flows which exchange heat with the other fluid (shell side) via the surface of the tubes themselves. The fluid therefore converges at the outlet of the tubes, in a chamber 22 which acts as a collector and which is formed inside the rear head 24, and is discharged from the heat exchanger 10 via a first outlet nozzle 26.

The fluid at the shell side is introduced into the heat exchanger 10 via a second inlet nozzle 28 and exchanges heat with the fluid which flows inside the tubes, passing over the outer surface of the tubes themselves via a constrained course. This constrained course is mainly imposed by the geometric shape of the shell 14, the geometric shape and the positioning of one or more suitable inner baffles 30 with respect to the shell 14 and the positioning of the tubes of the tube bundle 12. Once the travel in contact with the tubes is at an end, the fluid at the shell side is discharged from the shell 14 itself via a second outlet nozzle 32.

With reference to an extent direction of the tube bundle 12, the baffles 30 can be of the transverse type, as shown, for example, in FIG. 1, or of the longitudinal type, as shown, for example, in FIGS. 2A (two fluid passages at the shell side), 2B (divided fluid) and 2C (double divided flow). The material from which those baffles 30 are produced may be of the metal or plastics type. The transverse baffles are divided into plate type baffles (“plate baffles”, FIG. 1) and axial flow baffles (“axial flow baffles”), such as, for example, rod type baffles (“rod baffles”) which are shown in FIGS. 3A-3C. The plate baffles are the ones which are generally most used in order to increase the turbulence of the fluid and to obtain greater heat exchange coefficients. The axial flow baffles are used for applications where it is important to reduce the vibrations of the tubes owing to the flow.

The plate baffles may have a single segment with or without tubes in the window, may have multiple segments or may be in the form of a disk 30A and perforated disk 30B (“disk and doughnut baffles”), as shown in FIGS. 4A-4E. The single-segment baffles (FIGS. 4A and 4B) are obtained by cutting a portion, generally in the form of a circle segment, from a circular plate and are assembled inside the shell 14 in an alternating manner (180°) in order to produce a fluid form of the undulating type. That flow is quasi perpendicular to the tubes in the central zone contained between two adjacent baffles 30, while it is substantially axial (parallel with the tubes) in the zone crossing the window 36 (FIG. 1).

In the present specification, the term “window” 36 is intended to be understood to be the cross-section of passage of the fluid, which has a crossing direction parallel with the axis of the shell 14 and the tubes and which is delimited by the free edge 38 (also referred to as the “cutout” or non-constrained edge at the inner walls of the shell 14) of the respective baffle 30 at one side and by the inner profile of the shell 14, in the region of the intersection of that shell 14 with the plane of the baffle 30 itself, at the other side. The type of each baffle 30, the geometric shape thereof, the number of baffles 30 used and the positioning thereof inside the shell 14, as well as the geometric shape and the size of the cross-section of each window 36, have a relevant impact on the fluid-dynamic aspects of the fluid at the shell side in terms both of the coefficient of local heat exchange (and consequently also of mean total heat exchange over the entire exchange surface) and of total pressure drops.

In order to maximize the overall performance levels of heat exchange of a tube bundle 12 of which there are predetermined the diameter of the tubes, the number of tubes and the layout of the tubes (transverse “pitch” and longitudinal “pitch”), in the event that a shell 14 of circular cross-section is used, it results from the prior art that the best configuration is the one with baffles 30 with transverse plates with a single segment, without tubes in the window 36. That configuration ensures high heat exchange coefficients at the shell side with discrete homogeneity of operation for all the tubes because all the tubes exchange heat so that the flow of the fluid which passes over them at the shell side has a direction which is substantially orthogonal thereto (“cross-flow”). However, that circumstance is not produced in a configuration which provides for the presence of tubes in the window 36. In this second configuration, the tubes exchange heat in such a manner that the flow of the fluid which passes over it has a direction which is predominantly parallel with the tube itself (“parallel flow”), in particular in the zones adjacent to the window 36. This second heat exchange configuration generates exchange coefficients which are less than those of the first heat exchange configuration. A second advantage of the configuration without tubes in the window 36 involves a greater cross-section for the passage of the fluid through the window 36 itself, thereby ensuring smaller pressure drops for the same volumetric flow of fluid.

In the present specification, the term geometry of the crossing cross-sections of the tube bundle in “cross-flow” of the fluid at the shell side, between two adjacent baffles, is intended to be understood to be the effective area of passage of the secondary fluid through the tube bundle in the region of cross-sections which are orthogonal to the direction of flow and which are located in planes extending through the axes of the tubes of the same row. The area of passage is therefore net of the tubes. The term “row” is intended to be understood to be each group of tubes which are arranged in the same line.

In order to simplify the understanding of the drawings, shorten the specification and to be free of the need always to define a particular layout used for the tubes, or to be free from the definition both of a specific transverse “pitch”, which represents the distance between two axes of tubes which belong to the same row, and of a specific longitudinal “pitch”, which represents the distance between the plane which extends through the axes of a row of tubes and the parallel plane which extends through the axes of the tubes of the adjacent row (elements for which there exist different definitions in literature, but with the same meaning as the one provided here), the above-mentioned definition of geometry of the crossing cross-sections of the tube bundle in “cross-flow” of the fluid at the shell side will always be understood to be valid.

Similarly, the term “geometry” or “shape” of the crossing cross-section of the tube bundle in “cross-flow” refers in this instance not to the general effective geometric shape which results from the union of a plurality of small non-connected rectangular sections, but instead to a cross-section resulting from the intersection of the plane which extends through the axes of the tubes of a specific row of tubes with the surface of the shell (or of any inner partitions), and the planes of the two adjacent baffles. That section having simplified geometry makes it easier to describe and understand the differences between the solutions according to the prior art and the solutions introduced by the present invention.

One possible disadvantage of the configuration without tubes in the window 36 which could be established in some specific applications involves it being possible to use, once the dimensions of the shell 14 are fixed, a maximum number of tubes less than the maximum number which can be used in the configuration with tubes in the window 36. In some cases, though the total heat exchange coefficients are greater, for a fixed difference in logarithmic mean temperature between the two fluids, in the configuration without tubes in the window 36, as a result of the reduced heat exchange surface which can be used (smaller maximum number of tubes which can be used), the maximum total capacity of the heat exchanger 10 could also be less than that of the configuration with the tubes in the window 36. That limitation limits the preferable field of use of the configuration without tubes in the window 36, suggesting in general the use thereof in cases in which the cost connected with the tubes is the main cost item.

The present invention preferably but non-exclusively relates to dry expansion evaporators which are used in applications for air-conditioning and for refrigerating the process in the field of industrial refrigeration. In those applications, the main objective for the development of new heat exchange devices is to save energy in respect of the refrigerating units in which those devices are installed. In order to achieve that objective, attempts are made to maximize the efficiency of the heat exchange of the heat exchangers, mainly using tubes with exchange surfaces which are increased and optimized, the cost of which also contains a proportion which rewards the technological know-how, in terms of both application and production, in addition to the albeit important proportion depending on the material. In those applications, the costs connected directly and indirectly with the tubes are the main ones, for which one of the objectives of the development of new designs of heat exchangers, on the basis of a specific type of tube, is to increase the total heat exchange coefficient in order to be able to minimize the number of tubes for a specific thermal requirement.

A typical problem of dry expansion evaporators with a tube bundle and shell used in applications for air-conditioning and for refrigerating the process in the field of industrial refrigeration, which is strictly connected with the main objective of working with high levels of heat exchange efficiency, is to produce a homogeneous/uniform heat exchange condition for all the tubes, so that the refrigerating fluid which exchanges heat inside the tubes is discharged therefrom in a predetermined thermodynamic condition, independently of what the tube is. In particular, in the applications of refrigerating systems with screw type compressors, scroll type compressors or piston type compressors, for the refrigerating fluid which is discharged from the evaporator, in order to then be introduced into the inlet opening of the compressor, it is required that it be in the super-heated vapor state. Usually, the super-heating value is approximately 5 K (the value required measured in Kelvin usually varies from 3 K to 7 K), wherein the term super-heating value is intended to be understood to be the difference between the sensible temperature of the vapor and the saturation temperature at the pressure thereof at the measurement location. Maintaining this condition is fundamental for the reliability of the system (the objective is not to have a return of liquid refrigerant towards the inlet of the compressor, the presence of which could damage the compressor itself), for which reason this value of the degree of super-heating of the vapor of refrigerating fluid at the inlet to the compressor is the control parameter with which the entire unit is controlled by means of an expansion valve and a controller. In practice, it will be the case that the expansion valve controls the flow of refrigerating fluid which flows in the refrigerating circuit so as to obtain the desired super-heating value at the outlet from the evaporator.

Normally, there is produced a progression with oscillations of the controlled magnitude (the super-heating of the vapor) around the desired value, used as the “set-point” by the controller of the valve. In general, under operating conditions under which there are great oscillations around the set-point value of the controlled magnitude, there is produced a greater drop in efficiency of the heat exchange with respect to operating conditions with smaller oscillations. This fact results from the non-linearity of the phenomena of heat exchange around the desired working point.

One of the characteristics of the flow being discharged from the evaporator which may assist a specific control system (thermal expansion valve with a plurality of controllers) to work with oscillations having a reduced magnitude around the set-point value is the characteristic of homogeneity of the flow itself. A vapor flow at a non-homogeneous temperature, even with the presence of drops of liquid at the outlet from any tube, could in fact have a disruptive effect on the correct measurement of the temperature sensor, thereby initiating a so-called “pendulum” type operation of the system, characterized by oscillations of the super-heating with a great extent, with consequent losses of efficiency. One of the methods used to obtain the above-mentioned homogeneous operation for all the tubes is, in addition to the method of using an efficient system of distribution of the refrigerating fluid at the inlet of the tubes, the method of using a design of the heat exchanger at the shell side so as to produce a heat exchange homogeneity for all the tubes.

A second problem of dry expansion evaporators with a tube bundle and shell, which are always used in applications for air-conditioning and for refrigerating the process in the field of industrial refrigeration is that of the efficiency in the case of a multi-circuit configuration of the fluid at the tube side, in which the number of circuits is greater than two. In those applications, the heat exchanger configuration with a tube bundle and shell which is generally used is always the configuration with a cylindrical shell, with transverse plate baffles with a single segment positioned in such a manner, with respect to the positioning of the nozzles at the shell side, that the fluid at the shell side during the “cross-flow” flowing thereof through the tube bundle encounters in an alternating manner a plurality of circuits at the same time (see FIG. 5A, configuration with three circuits, and FIG. 5B, configuration with four circuits).

The circuits are configured with a non-symmetrical geometry because it is sought to arrange the different circuits with such a geometry as to allow the use of refrigerating fluid discharge connections which are sufficiently large not to bring about great losses of efficiency. Those losses of efficiency result from the pressure drops of the fluid during the crossing of those discharge connections. Those losses of efficiency become higher with the refrigerating fluid used becoming more sensitive to the pressure drops. The limit on the dimensions of those discharge connections is imposed by the difficulty of assembly as a result of the physical interferences between the connections of the different circuits.

The use of similar or identical circuits, with a symmetrical geometry, which are crossed with a “cross-flow” by the fluid at the shell side in a symmetrical and uniform manner, that is to say, in such a manner that each section of the flow at the shell side with a uniform velocity exchanges heat in a homogeneous manner with the tubes of a single circuit each time, would require the use of discharge connections of the refrigerating fluid having smaller dimensions as a result of the geometry itself of the circuits (see FIG. 6). The solution which is generally used (that of FIGS. 5A and 5B) involves the problem of the substantial loss of efficiency in the case of operation with a partial charge, that is to say, when one or more circuits are deactivated. In those cases, the flow of the fluid at the shell side exchanges heat in a non-symmetrical/non-uniform manner with the portion of the surface of the active heat exchange, thereby generating a loss of efficiency.

Another problem of the multi-circuit applications at the tube side involves the fact that sometimes, in order to allow the use of discharge connections having greater dimensions for the refrigerating fluid, the use of a configuration with transverse plate baffles with a single segment with tubes in the windows is selected (see FIG. 7). This solution, in addition to having negative effects both with respect to the total exchange coefficients thereof and with respect to the pressure drops at the shell side with respect to a solution without tubes in the window, also adds an additional non-uniformity of operation between the various circuits. Some circuits (the innermost circuits) in fact could be working in a purely “cross-flow” condition at the shell side, while other circuits could work in a mixed condition (“cross-flow”/“parallel flow”) with lesser efficiency. This impairment would also be produced in the case of identical circuits with symmetrical geometry (see FIG. 8).

Heat exchangers with a tube bundle and shell which are provided with circular baffles are known from the prior art. The most common embodiments use a cylindrical shell with transverse plate baffles with a single segment, with or without tubes in the window. The form of the window of the baffle which has been determined is that of a circle segment. Similar embodiments are illustrated in FIGS. 5A, 5B, 6, 7 and 8.

The document WO98/08031 A1 sets out a heat exchanger which is provided with circular baffles with a single segment with a double square cutout. In fact, that heat exchanger is provided with a shell having a circular cross-section, with inner squaring of the cross-section of the tube bundle and with a window mainly in the form of a circle segment. The technical solution adopted in the document WO98/08031 A1 provides for the insertion of two partitions inside the cylindrical shell in order to homogenize the behavior of the tubes. By this solution being adopted, the geometry of the crossing cross-sections of the tube bundle in “cross-flow”, between two adjacent baffles, is constant and is of rectangular or square shape. However, the baffle, though not described in document WO98/08031 A1, will presumably have a circular geometry with a single segment, with two “squared” cutouts in the region of the partitions which are positioned inside the shell, so as to limit the bypasses. The window of the baffle remains mainly in the form of a circle segment.

The document CN104625864 A sets out a heat exchanger whose shell may also have a square shape in cross-section. Consequently, the respective baffles may have a squared geometry of rectangular shape while the window has a rectangular shape and the tubes are inside the window itself. That configuration with a square shell is proposed in order to allow a solution to the problem of cleaning the tube bundle. Similar embodiments to that of the document CN104625864 A are described in the additional documents CN202041098 U and CN203359975 U.

SUMMARY OF THE INVENTION

A general object of the present invention is therefore to provide a heat exchanger with a tube bundle and shell, particularly but not exclusively a dry expansion evaporator for air-conditioning and for refrigerating a process in the field of industrial refrigeration, which is capable of overcoming the above-cited disadvantages of the prior art in a manner which is extremely simple, economical and particularly functional.

In detail, an object of the present invention is to provide a heat exchanger with a tube bundle and shell which is capable of carrying out heat exchange processes with a high level of efficiency so as to minimize the quantity of tubes used.

Another object of the invention is to provide a heat exchanger with a tube bundle and shell which is capable of carrying out homogeneous/uniform heat exchange processes for all the tubes so as to obtain super-heating of the vapor which is as homogeneous/uniform at the discharge of the tubes themselves, thereby minimizing the extents of the oscillations during the control of the heat exchanger and consequently increasing the efficiency.

Another object of the invention is to provide a heat exchanger with a tube bundle and shell which is capable of containing the value of the pressure drops at the shell side, particularly in the case of a high level of filling of tubes. Those pressure drops are normally produced both owing to the numerous “cross-flow” passages and owing to the limitations in the free surface for the passage of the fluid during crossing of the window of the baffle (with or without the tubes in the window).

Yet another object of the invention is to construct a heat exchanger with a tube bundle and shell in which, in the multi-circuit applications at the tube side, the functionality of the circuits is increased under all charging conditions with discharge connections of the refrigerating fluid having such dimensions as not to introduce excessive pressure drops which would have consequent negative effects on the efficiency.

These objects according to the present invention are achieved by providing a heat exchanger with a tube bundle and shell as set out in claim 1.

Additional features of the invention are set out in the dependent claims which are an integral part of the present description.

Generally, the fluid-dynamic configuration at the shell side of a heat exchanger with a tube bundle and shell according to the present invention is characterized as follows:

geometry of the crossing cross-sections of the tube bundle in “cross-flow” of the fluid between two adjacent baffles being constant and having a rectangular or square shape;

configuration with plate type baffles having a rectangular or square geometry;

rectangular geometry of the window for crossing the baffle, without any tubes in the window.

The characteristics and advantages of a heat exchanger with a tube bundle and shell according to the present invention will be better appreciated from the following exemplary and non-limiting description, with reference to the appended schematic drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIGS. 1A and 1B are cross-sections which show the main components of a heat exchanger with a tube bundle and shell in the configuration with a single passage and in the configuration with multiple passages (double) for the fluid at the tube side, respectively,

FIGS. 2A, 2B and 2C show respective configurations of heat exchangers with a tube bundle and shell with baffles of the longitudinal type;

FIG. 3A shows a set of baffles with axial flow through a heat exchanger with a tube bundle and shell;

FIGS. 3B and 3C show some specific configurations of the axial flow baffles of FIG. 3A;

FIGS. 4A and 4B show two configurations of the baffles in a heat exchanger with a tube bundle and shell, with tubes in the window and without tubes in the window, respectively;

FIG. 4C shows a configuration of a baffle with a double segment in a heat exchanger with a tube bundle and shell;

FIG. 4D shows a configuration of a baffle with a triple segment in a heat exchanger with a tube bundle and shell;

FIG. 4E shows a configuration of a baffle with a disk and perforated disk in a heat exchanger with a tube bundle and shell;

FIG. 5A shows a heat exchanger with a tube bundle and shell in a configuration with three circuits for the fluid at the shell side;

FIG. 5B shows a heat exchanger with a tube bundle and shell in a configuration with four circuits for the fluid at the shell side;

FIG. 6 shows a heat exchanger with a tube bundle and shell in another configuration with three circuits for the fluid at the shell side, without tubes in the window;

FIG. 7 shows a heat exchanger with a tube bundle and shell in another configuration with four circuits for the fluid at the shell side, without tubes in the window;

FIG. 8 shows a heat exchanger with a tube bundle and shell in another configuration with three circuits for the fluid at the shell side, with tubes in the window;

FIG. 9A is a perspective view of a first type of heat exchanger with a tube bundle and shell according to the prior art;

FIG. 9B shows a qualitative profile of the component in “cross-flow” of the speed vector of the fluid at the shell side, obtained in the region of the “cross-section A” and the “cross-section B” of FIG. 9A, respectively;

FIG. 10A is a perspective view of another type of heat exchanger with a tube bundle and shell according to the prior art;

FIG. 10B shows a qualitative profile of the component in “cross-flow” of the speed vector of the fluid at the shell side, obtained in the region of the “cross-section A” and the “cross-section B” of FIG. 10A, respectively;

FIG. 11A is a perspective view of a heat exchanger with a tube bundle and shell according to the present invention;

FIG. 11B shows a qualitative profile of the component in “cross-flow” of the speed vector of the fluid at the shell side, obtained in the region of the “cross-section A” and the “cross-section B” of FIG. 11A, respectively;

FIG. 12 is a cross-section of a first embodiment of the heat exchanger with a tube bundle and shell according to the present invention, with a shell having a square cross-section;

FIG. 13 is a cross-section of a second embodiment of the heat exchanger with a tube bundle and shell according to the present invention, with a shell having a rectangular cross-section;

FIG. 14 is a cross-section of a third embodiment of the heat exchanger with a tube bundle and shell according to the present invention, with a shell having a circular cross-section and a square baffle; and

FIG. 15 is a cross-section of a fourth embodiment of the heat exchanger with a tube bundle and shell according to the present invention, with a shell having a circular cross-section and a rectangular baffle.

DETAILED DESCRIPTION OF THE INVENTION

It should be noted that, in the appended Figures and in the following description, identical reference numerals indicate mutually identical or equivalent elements.

With particular reference to FIGS. 11A to 15, there are shown some preferred embodiments of the heat exchanger with a tube bundle and shell according to the present invention, which heat exchanger is generally designated 10. The heat exchanger 10 comprises in a manner known per se a plurality of tubes 12 which are arranged parallel with each other in order to form one or more tube bundles. The tubes 12 are inserted axially into a member 14 which is of elongate form and which has a cylindrical geometry and which constitutes the shell of the heat exchanger 10. In the present specification, the term “member having a cylindrical geometry” is intended to be understood to be the locus of the generating lines parallel with a predefined direction and passing through the single points of a predetermined directrix curve and which is delimited by two opposing bases which are substantially parallel with the directrix. The directrix, which is always a closed line, may be, for example, a square (FIG. 12), a rectangle (FIG. 13), a circle (FIGS. 14 and 15), or may generally be in the form of any polygon.

A first fluid, fed through one or more first inlet nozzles 16, placed at a first end of the shell 14 and axially orientated, is capable of flowing inside the tubes 12 of the tube bundle and is discharged via a first discharge nozzle 26, positioned in the region of the opposite end of the shell 14 and also axially orientated. A second fluid, fed through at least one second inlet nozzle 28, typically positioned on the side surface of the shell 14, instead flows inside the shell 14 itself and flows over the outer walls of the tubes 12. The second fluid is discharged via a second discharge nozzle 32, also positioned on the side surface of the shell 14. Therefore, the heat exchange between the first fluid and the second fluid is brought about via the walls of the tubes 12.

In the configuration which has a single passage for the fluid at the tube side and which is shown in FIG. 1A, the opposite ends of the tubes 12 of the tube bundle are connected to two tube plates 34 which are positioned in the region of the first inlet nozzle 16 at one side and the first outlet nozzle 26 at the other side. The tube plates 34 allow separation of the second fluid at the shell side, that is to say, the fluid which flows inside the shell 14, from the first fluid at the tube side, that is to say, the fluid which flows inside the tubes 12 of the tube bundle.

In the configuration which has a double passage for the fluid at the tube side and which is shown in FIG. 1B, however, it is possible to use tubes 12 which are U-shaped or with a different suitable geometry, wherein the open ends of each tube 12 are fixed to the same tube plate 34 which is arranged in the region of an end of the shell 14, while in the region of the opposite end of this shell 14 there is provided a closure base 42 which is not perforated.

Apart from the configuration of the heat exchanger 10, there are formed inside the shell 14 two or more baffles 30 which are arranged perpendicularly with respect to the centre axis of the shell 14 itself. There is defined between each baffle 30 and the inner walls of the shell 14 a corresponding window 36 which is constituted by the cross-section for the passage of the second fluid, with a crossing direction parallel with the centre axis of the shell 14 and the tubes 12, delimited by the free edge 38 of the respective baffle 30 on the one side and by the inner profile of the shell 14, at the intersection of this shell 14 with the plane of the baffle 30 itself, on the other side.

According to the invention, the cross-section for the passage of the second fluid which is positioned between two adjacent baffles 30 is constant and has a rectangular shape, that is to say, with all the inner angles congruent with each other. In particular, the cross-section for the passage of the second fluid which is positioned between two adjacent baffles 30 may have a square shape, that is to say, both with all the inner angles and with the four sides congruent with each other. In addition, each baffle 30 also has a rectangular shape or more particularly a square shape. Finally, each window 36 also has a rectangular shape and does not have any tubes 12 therein.

The above-mentioned configurations may be constructed by using a shell 14 with a cross-section, at its own longitudinal axis, of square shape (FIG. 12) or of rectangular shape (FIG. 13), or by using suitable longitudinal partitions 40 which are positioned inside a shell 14 with a cross-section having a circular shape (FIGS. 14 and 15). In this last case, the longitudinal partitions 40 form the walls that surround the windows 36 and generally the cross-sections for the passage of the second fluid inside the heat exchanger 10.

It has thereby been seen that the heat exchanger with a tube bundle and shell according to the present invention achieves the objectives set out above. Using the configuration to which the invention relates, the flow of fluid at the shell side will be confined to the heat exchanger with the tube bundle via cross-sections of passage having a constant area for each crossing between two adjacent baffles, thereby benefitting from a greater uniformity including along the axis which is orthogonal to the direction of flow during the crossing in “cross-flow”, as a result of the use of windows with a geometry with double symmetry, which are rectangular or square, without tubes extending therein.

With respect to the heat exchangers of the known type, with a cylindrical shell having a circular directrix curve with plate baffles with a single segment and with or without tubes in the window, the heat exchanger with a tube bundle and shell according to the present invention allows the following advantages to be achieved:

greater heat exchange efficiency and consequently a reduction in the exchange surface necessary;

homogeneous/uniform operation of all the tubes, with resultant greater stability during control in applications with a refrigerating cycle;

greater heat exchange efficiency in applications with a multi-circuit (more than two circuits) at the tube side both at full charge (that is to say, with all the circuits active) and at partial charge (for example, with a single active circuit).

Those advantages are achieved as a result of:

constant geometry of the cross-sections of passage of the tube bundle in “cross-flow” of the fluid between two adjacent baffles (see FIG. 11A);

geometry with double symmetry (rectangular or square) of the window of the baffle, without tubes extending therein, which allows greater uniformity of heat exchange (better redistribution of the fluid after crossing through the window) including along the axis which is orthogonal to the direction of flow during the crossing in “cross-flow” (see FIG. 11B);

identical geometry and identical conditions of heat exchange for each circuit, with consequent homogeneous/uniform operation of all the tubes of the active circuits;

geometry of each circuit which allows the use of connections for the refrigerating fluid having optimized dimensions in order to reduce the pressure drops.

The heat exchanger with a tube bundle and shell configured in this manner is capable in any case of a number of modifications and variants, all included within the same innovative concept; furthermore, all the details can be replaced with technically equivalent elements. In practice, the materials used as well as the forms and the dimensions may be freely selected in accordance with the technical requirements.

The scope of teaching of the invention is therefore defined by the appended claims. 

1. Heat exchanger comprising a plurality of tubes arranged parallel to each other in order to form one or more tube bundles axially inserted into a body which has an elongated shape and a cylindrical geometry forming the shell of the heat exchanger, a first fluid, fed through one or more first inlet nozzles, placed at a first end of the shell and axially oriented, flowing inside the tubes and a second fluid, fed through at least one second inlet nozzle, flowing inside said shell in order to perform the heat exchange with the first fluid through the walls of the tubes, inside the shell two or more baffles being obtained, arranged perpendicularly with respect to the central axis of said shell, between each baffle and the inner walls of the shell a corresponding window being defined, consisting of the cross-section for the passage of the second fluid, with a crossing direction parallel to the central axis of the shell and to the tubes, delimited by the free edge of the respective baffle on the one side and by the inner profile of the shell, at the intersection of said shell with the plane of said baffle, on the other side, in the heat exchanger the cross-section for the passage of the second fluid placed between two adjacent baffles is constant and has a rectangular shape, i.e. all its inner angles are congruent to each other, each baffle has a rectangular shape, each window has a rectangular shape and has no tubes therein.
 2. Heat exchanger according to claim 1, wherein the cross-section for the passage of the second fluid placed between two adjacent baffles has a square shape, i.e. all its inner angles and its four sides are congruent to each other.
 3. Heat exchanger according to claim 1, wherein each baffle has a square shape.
 4. Heat exchanger according to claim 1, wherein the shell has a rectangular-shaped cross-section at its longitudinal axis.
 5. Heat exchanger according to claim 4, wherein the shell has a square-shaped cross-section at its longitudinal axis.
 6. Heat exchanger according to claim 1, wherein the shell has a circular-shaped cross-section and is internally provided with longitudinal partitions, said longitudinal partitions forming the walls that surround the windows and the cross-sections for the passage of the second fluid inside the heat exchanger.
 7. Heat exchanger according to claim 1, wherein it comprises straight tubes whose opposed open ends are connected with two respective tube plates, said tube plates separating the second fluid from the first fluid.
 8. Heat exchanger according to claim 1, wherein it comprises tubes whose open ends are fixed at a same tube plate arranged at an end of the shell, whereas a non-perforated closing bottom is provided at the opposite end of said shell.
 9. Heat exchanger according to claim 1, wherein the second inlet nozzle is placed on the side surface of the shell. 